Automatic transmission control system for automobiles

ABSTRACT

An automatic transmission control system for an automotive automatic transmission including a speed sensor for detecting a vehicle speed of the vehicle, a load sensor for detecting a load acting on the driving torque generator, a torque sensor for detecting input torque transmitted to the automatic transmission from the driving torque generator, first working fluid pressure control communicating both a first friction coupling element&#39;s unlocking pressure chamber and a second friction coupling element&#39;s pressure chamber with the working fluid source in the hydraulic pressure control system for controlling supply of working fluid pressure to and from both the unlocking pressure chambers and a second working fluid pressure control in the fluid path communicating both the first friction coupling element&#39;s locking pressure chamber with the working fluid source for controlling supply of working fluid pressure to and from the first friction coupling element&#39;s locking pressure chamber. Thereby controlling the second working fluid pressure control such that, until a start of a substantial gear shift period of time from an appearance of the specific gear shift command signal, the level of the working fluid pressure supplied to the locking pressure chamber of the first friction coupling element being directly related to a level of the working fluid pressure supplied to the locking pressure chamber of the first friction coupling element during the substantial gear shift period of time.

BACKGROUND OF THE INVENTION

1. Field of the Invention

This invention relates to an automatic transmission control system foran automobile, and more particularly, to an improvement of an automatictransmission control system in which shift control for a specific gearshift is performed by especially controlling a hydraulic workingpressure.

2. Description of the Related Art

Automatic transmissions for automotive vehicles, which typicallycomprise a torque converter and a transmission gear mechanism, areautomatically shifted into desired gears by selectively locking andunlocking a plurality of friction coupling elements such as clutches andbrakes to switch the torque transmission path of the transmission gearmechanism. Such an automatic transmission is provided with hydraulicpressure control systems which supplies a pressurized working fluid toand discharges a pressurized working fluid from the respective frictioncoupling elements to selectively lock and unlock them. This type ofautomatic transmission includes, in addition to regular types offriction coupling elements, a servo cylinder type of friction couplingelement provided with a band brake. This servo cylinder type frictioncoupling element has a servo cylinder which is divided into two pressurechambers, namely a servo apply pressure chamber and a servo releasepressure chamber, by means of a spring loaded piston ordinarily forcedin such a direction as to move toward the servo release pressurechamber. The servo cylinder type friction coupling element is lockedwhen only the servo apply pressure chamber is supplied with apressurized working fluid, and unlocked when both servo apply pressurechamber and servo release pressure chamber are not supplied with apressurized working fluid, when both servo apply pressure chamber andservo release pressure chamber are supplied with a pressurized workingfluid, and when only the servo release pressure chamber is supplied witha pressurized working fluid.

Some of available gear shifts have a necessity of locking a specificfriction coupling element simultaneously with unlocking another specificfriction coupling element. For example, in the case where the servocylinder type friction coupling element is employed as a 2-4 brake to belocked in a second gear and in a fourth gear, and a single pressurechamber friction coupling element is employed as a 3-4 clutch to belocked in a third gear and in the fourth gear, it is necessary for a 2-3gear shift to lock the 3-4 clutch simultaneously with unlocking the 2-4brake. In this instance, because the pressurized working fluid issimultaneously supplied to both 2-4 brake and 3-4 clutch, the hydraulicpressure control circuit has fluid paths which branch off from a commonfluid path and lead to the pressure chambers of these 2-4 brake and 3-4clutch, respectively and is provided with a single hydraulic pressureregulating valve, such as a duty solenoid valve, located at the junctionbetween the branch fluid paths and common fluid path in order to controlsupply of the pressurized working fluid to the pressure chambers of the2-4 brake and 3-4 clutch. Such a hydraulic pressure control system foran automotive automatic transmission is known from, for example,Japanese Patent Publication No. 6-21643.

In the prior art hydraulic pressure control system, while the 3-4 clutchis supplied with a pressurized working fluid (which is referred to as aclutch locking fluid pressure), the 2-4 brake at the servo releasepressure chamber is simultaneously supplied with a pressurized workingfluid (which is referred to as a servo releasing fluid pressure) toforce the piston in such a direction as to release the 2-4 brake. It ishard for the hydraulic pressure control system to perform precisecontrol of the pressurized working fluid during a movement of thepiston, and hence it is impossible to perform precise gear shift controlwhich causes the turbine speed to fall keeping the change rate ofturbine speed in agreement with a target rate through the control of 3-4clutch locking pressure during a gear shift.

In regard to this, it may be done to control the 3-4 clutch lockingfluid pressure indirectly through the servo releasing fluid pressure bycontrolling a pressurized working fluid in the servo apply pressurechamber of the 2-4 brake (which is referred to as a servo applying fluidpressure). In this case, even while the piston moves in the servocylinder of the 2-4 brake, it is enabled to control the 3-4 clutchlocking fluid pressure so as to cause, for example, an appropriatechange in turbine speed during a gear shift. When controlling the 3-4clutch locking fluid pressure indirectly by means of the servo applyingfluid pressure, the servo applying fluid pressure, which is at arelatively high pressure level and has been supplied to provide a secondgear, is fallen once with an effect of causing smooth control of the 3-4clutch locking fluid pressure, and kept at the fallen pressure level toprovide a torque phase. Afterward, when the gear shift is actuallyactivated, in other words when the gear shift control enters on aninertia phase during which the turbine speed falls, the control of servoapplying fluid pressure is performed so as to cause a satisfactory 3-4clutch locking action.

During this control of servo applying fluid pressure, a serious problemis encountered. Specifically, when the servo applying fluid pressure isfallen for the purpose of controlling the 3-4 clutch locking fluidpressure, it is set to a pressure level suitable for the 3-4 clutch tobe smoothly locked which is a specific pressure level meeting to, forexample, a turbine speed change rate yielded due to input torque to the3-4 clutch and a locking action of the 3-4 clutch. In some cases, thespecific pressure level is insufficient against input torque to the 2-4brake, allowing slippage of the 2-4 brake in advance of initiation ofthe 3-4 clutch locking action. This advanced slippage of the 2-4 brake,which is significant especially in the case where, while the specificpressure level of 3-4 clutch locking fluid pressure is made low due to alow turbine speed change rate when the engine is operating in a range oflower speeds, the 2-4 brake input torque increases more than thespecific pressure level of 3-4 clutch locking fluid pressure increasesin accordance with engine loads as a result of having entered anoperating range of higher engine loads, yields an engine blow.

SUMMARY OF THE INVENTION

It is an objective of the present invention to provide an automatictransmission control system for an automotive automatic transmissionwhich causes satisfactorily a specific gear shift, such as a 2-3 gearshift which has a necessity of locking a specific friction couplingelement simultaneously with unlocking another specific friction couplingelement.

The foregoing object of the present invention is achieved by providing acontrol system for an automotive automatic transmission comprised of atransmission gear mechanism equipped with a plurality of frictioncoupling elements, and a hydraulic pressure control circuit whichsupplies a pressurized fluid to and discharges a pressurized fluid fromthe friction coupling elements to selectively lock and unlock them tochange the torque transmission path in the transmission gear mechanismand thereby to provide desired gears. The friction coupling elementsinclude a first and second specific friction coupling elements lockedand unlocked to provide a specific gear. The first friction couplingelement is of a cylinder type which has a cylinder chamber divided intoa servo apply pressure chamber and a servo release pressure chamber by apiston and is locked when only the servo apply pressure chamber issupplied with a pressurized working fluid and unlocked when at least theservo release pressure chamber is supplied with a pressurized workingfluid and when both servo apply pressure chamber and servo releasepressure chamber are not supplied with a pressurized working fluid. Thesecond specific friction coupling element has a single pressure chamberand is locked when the pressure chamber is supplied with a pressurizedworking fluid. The second friction coupling element is in communicationwith the servo release pressure chamber of the first specific frictioncoupling element. The control system has first and second fluid controlmeans according to driving conditions to automatically cause a gearshift meeting to the driving condition. Specifically, during a specificgear shift which needs a change from a state where, while the servoapply pressure chamber of the first specific friction coupling elementis supplied with a pressurized working fluid, both servo apply pressurechamber and pressure chamber are not supplied with a pressurized workingfluid to a state where all of the servo apply pressure chamber, theservo release pressure chamber and the pressure chamber are suppliedwith a pressurized working fluid, the control system controls the secondfluid control means to make the pressurized working fluid supplied tothe servo apply pressure chamber lower in pressure level during a gearshift period for which a change in rotational speed input to thetransmission gear mechanism is caused due to the specific gear shiftthan before and after the gear shift period, to cause the pressurizedworking fluid supplied to the servo apply pressure chamber to reachclosely to a pressure level attained in the gear shift period within aprevious period from an appearance of a shift command for the specificgear shift to the beginning of the gear shift period, and, according tothe input torque, to prevent the pressurized working fluid supplied tothe servo apply pressure chamber during the previous period fromdropping a critical pressure level. This critical pressure level may bespecified as a level at which the transmission gear mechanism at itsrotational speed input means is allowed to run idle, or otherwise as alevel determined in accordance with the input torque.

According to another embodiment, the automatic transmission controlsystem controls the second fluid control means to make the pressurizedworking fluid supplied to the servo apply pressure chamber to becomeequal to a pressure level attained in the gear shift period within theprevious period.

With the automatic transmission control system thus structured, in thecase where, during the specific gear shift which is caused by unlockingthe first specific friction coupling element and simultaneously lockingthe second specific friction coupling element, the pressurized workingpressure of the second specific friction coupling element during thespecific gear shift is controlled indirectly by control of thepressurized working pressure of the servo apply pressure chamber of thefirst specific friction coupling element through the pressurized workingpressure of the servo release pressure chamber of the first specificfriction coupling element, because the pressurized working pressure ofthe servo apply pressure chamber of the first specific friction couplingelement is changed to be close to the pressure level attained during thespecific gear shift before the commencement of actual shift operation,the control of pressurized working pressure of the second specificfriction coupling element during the actual shift operation is initiatedsmoothly. Furthermore, when the pressurized working pressure of theservo apply pressure chamber of the first specific friction couplingelement is changed, it is prevented from dropping below the criticalpressure level, so that an engine blow-up operation due to slippage ofthe incompletely locked first specific friction coupling element. Forexample, even in the case where the pressurized working pressure of theservo apply pressure chamber of the first specific friction couplingelement, after having been changed, is set at such a pressure level asto initiate smoothly locking the second specific friction couplingelement, and as a result, the pressurized working pressure of the servoapply pressure chamber of the first specific friction coupling elementis expected to drop below a pressure level necessary for the firstspecific friction coupling element to remain locked in, in particular, arange of low engine speeds and high engine loads, it is effectivelyprevented from encountering such an undesirable drop. The pressurizedworking fluid supplied to the servo apply pressure chamber of the firstspecific friction coupling element is changed to be equal to a pressurelevel attained in the gear shift period within the previous period.Furthermore, the critical pressure level is specified as a level atwhich the rotational speed input shaft of transmission gear mechanism isallowed to run idle, or otherwise as a level determined in accordancewith the input torque, the gear shift is certainly prevented fromstarting with a delay due to over restriction of a drop in thepressurized working fluid supplied to the servo apply pressure chamberof the first specific friction coupling element when the input torque islow, or from yielding an engine blow-up operation due to an insufficientdrop in the pressurized working fluid supplied to the servo applypressure chamber of the first specific friction coupling element whenthe input torque is high.

BRIEF DESCRIPTION OF THE DRAWINGS

The foregoing and other objects and features of the present inventionwill be clearly understood from the following detailed description ofpreferred embodiments when read in conjunction with the accompanyingdrawings, in which:

FIG. 1 is a schematic skeleton view showing a mechanical structure of anautomatic transmission equipped with a control system of the presentinvention;

FIG. 2 is a cross-sectional view of a transmission gear mechanism of theautomatic transmission shown in FIG. 1;

FIG. 3 is a hydraulic control circuit of an automatic transmissioncontrol system in accordance with an embodiment of the presentinvention;

FIG. 4 is a cross-sectional view of a hydraulic actuator for a 2-4brake;

FIG. 5 is a block diagram illustrating a control system of varioussolenoid valves installed in the hydraulic control circuit;

FIG. 6 is an enlarged view of an essential part of the hydraulic controlcircuit of FIG. 3 which is in a state for a first gear;

FIG. 7 is an enlarged view of an essential part of the hydraulic controlcircuit of FIG. 3 which is in a state for a second gear;

FIG. 8 is an enlarged view of an essential part of the hydraulic controlcircuit of FIG. 3 which is in a state for a third gear;

FIG. 9 is an enlarged view of an essential part of the hydraulic controlcircuit of FIG. 3 which is in a state for a forth gear;

FIG. 10 is an enlarged view of an essential part of the hydrauliccontrol circuit of FIG. 3 which is in a state for the first gear in alow (L) range;

FIG. 11 is an enlarged view of an essential part of the hydrauliccontrol circuit of FIG. 3 which is in a state for a reverse gear

FIG. 12 is an explanatory diagram of turbine speed change rate withrespect to time during an up shift;

FIG. 13 is an explanatory diagram of transmission output torque duringan up shift;

FIG. 14 is a flowchart illustrating the sequence routine of first dutysolenoid control during a 1-2 gear shift;

FIG. 15 is a time chart illustrating controlling and controlledparameters during the 1-2 gear shift;

FIG. 16 is a flowchart illustrating the sequence routine of basehydraulic pressure calculation performed during the 1-2 gear shift;

FIG. 17 is a map of hydraulic pressure with respect to target turbinespeed change rate used in the base hydraulic pressure calculation;

FIG. 18 is a map of hydraulic pressure with respect to target turbinetorque used in the base hydraulic pressure calculation;

FIG. 19 is a map of a square of hydraulic pressure with respect totarget turbine torque used in the base hydraulic pressure calculation;

FIG. 20 is a flowchart illustrating the sequence routine of feedbackhydraulic pressure calculation during the 1-2 gear shift;

FIG. 21 is an explanatory diagram of an example of membership functionsused in the feedback hydraulic pressure calculation;

FIG. 22 is an explanatory diagram of another example of membershipfunctions used in the feedback hydraulic pressure calculation;

FIG. 23 is an explanatory diagram of another example of membershipfunctions used in the feedback hydraulic pressure calculation;

FIG. 24 is an explanatory diagram of fuzzy stratified zones;

FIG. 25 is a flowchart illustrating the sequence routine of learnedpressure calculation during the 1-2 gear shift;

FIG. 26 is a time chart illustrating controlling and controlledparameters during the learned pressure calculation;

FIG. 27 is an explanatory diagram of an example of membership functionson peak values of turbine speed change rates used in the learnedpressure calculation;

FIG. 28 is an explanatory diagram of an example of membership functionson average values of feedback-manipulated variables used in the learnedpressure calculation;

FIG. 29 is an explanatory diagram of an example of membership functionson average values of deviations of feedback-manipulated variables usedin the learned pressure calculation;

FIG. 30 is a flowchart illustrating the sequence routine of initialworking pressure correction control during the 1-2 gear shift;

FIG. 31 is a time chart illustrating parameters of the initial workingpressure correction control;

FIG. 32 is a flowchart illustrating the sequence routine of prechargecontrol during the 1-2 gear shift;

FIG. 33 is a map of flowing quantity of a base pressurized fluid withrespect to line pressure used in the precharge control;

FIG. 34 is a map of control coefficient with respect to fluidtemperature used in the precharge control;

FIG. 35 is a flowchart illustrating the sequence routine of first dutysolenoid control during a 2-3 gear shift;

FIG. 36 is a map of lower limit hydraulic pressure with respect toturbine torque used during the 2-3 gear shift;

FIG. 37 is a flowchart illustrating the sequence routine of second dutysolenoid valve control during the 2-3 gear shift;

FIG. 38 is a time chart illustrating controlling and controlledparameters during the 2-3 gear shift;

FIG. 39 is a time chart illustrating another example of servo applypressure control during the 2-3 gear shift;

FIG. 40 is a time chart illustrating still another example of servoapply pressure control during the 2-3 gear shift;

FIG. 41 is a time chart illustrating controlling and controlledparameters in the precharge control during the 2-3 gear shift;

FIG. 42 is a map of initial count with respect to engine throttleopening used in the precharge control;

FIG. 43 is an explanatory diagram of transmission output torque duringthe precharge control when transmission input torque is low; and

FIG. 44 is a time chart illustrating controlling and controlledparameters during the precharge control in the 2-3 gear shift.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT Mechanical Structure

Referring to the drawings in detail, in particular, to FIG. 1 which is askeleton diagram schematically showing the entire mechanical structureof an automatic transmission 10 in which an automatic transmissioncontrol system in accordance with an embodiment of the present inventionis incorporated, the automatic transmission 10 has, as its mainstructural elements, a hydraulic torque converter 20 and front and rearplanetary gear mechanisms 30 and 40 arranged contiguously to each otheras a transmission gear mechanism which are driven by means of outputtorque of the torque converter 20. The automatic transmission 10 furtherhas a plurality of friction coupling elements 51-55, such as clutches,brakes and so forth, and a one-way clutch 56 which are selectivelylocked and unlocked to switch the power transmission path of the frontand rear planetary gear mechanism 30 and 40, placing the automatictransmission 10 into desired gears, namely first (1st) to fourth (4th)gears in a drive (D) range, first (1st) to third (3rd) gears in a second(S) range, first (1st) and second (2nd) gears in a low (L) range, and areverse (RV) gear in a reverse (R) range.

The hydraulic torque converter 20 is comprised of a pump 22 locatedwithin a converter housing 21 which is fastened to an engine outputshaft 1, a turbine 23 which is arranged to face to the pump 22 anddriven by the pump 22 through a working fluid, and a stator 25 which issupported between the pump 22 and the turbine 23 by a transmissionhousing 11 through a one-way clutch 24 and does multiply engine outputtorque. The hydraulic torque converter 20 is provided with a lockupclutch 26 between the converter housing 21 and the turbine 23 tomechanically lock the engine output shaft 1 (pump 22) and the turbine 23together when the lockup clutch 26 is activated. Transmission of enginetorque is made from the turbine 23 to the front and rear planetary gearmechanisms 30 and 40 through a turbine shaft 27 fastened to the turbine23. An oil pump 12, which is driven by the engine output shaft 1 throughthe converter housing 21 of the hydraulic torque converter 20, isarranged on one side of the hydraulic torque converter 20 opposite tothe engine.

Each of the front and rear planetary gear mechanisms 30 and 40 iscomprised of a sun gear 31, 41, and a plurality of pinions 32, 42 inmesh with the sun gear 31, 41, a pinion carrier 33, 43 which supportsthese pinions 32, 42, and a ring gear 34, 44 in mesh with the pinions32, 42. There are provided in the transmission gear mechanism a forwardclutch (FWCL) 51 between the turbine shaft 27 and the sun gear 31 of thefront planetary gear mechanism 30, a reverse clutch (RVCL) 52 betweenthe turbine shaft 27 and the sun gear 41 of the rear planetary gearmechanism 40, a 3rd-4th clutch (3-4CL) 53 between the turbine shaft 27and the pinion carrier 43 of the rear planetary gear mechanism 40, and a2nd-4th (2-4) brake (2-4BR) 54 which locks the sun gear 41 of the rearplanetary gear mechanism 40. Between these front and rear planetary gearmechanisms 30 and 40, the pinion carrier 33 and the ring gear 34 of thefront planetary gear mechanism 30 are linked with the ring gear 44 andthe pinion carrier 43 of the rear planetary gear mechanism 40,respectively. A low-reverse brake (LRBR) 55 and the one-way clutch(OWCL) 56 are arranged in parallel with respect to these pinion carrier33 and ring gear 44 and interposed between these pinion carrier 33 andring gear 44 and the transmission housing 11. Further, there is providedin the transmission gear mechanism an output gear 13 in mesh with thepinion carrier 33.

An intermediate transmission mechanism 60 includes a front intermediategear 62 fastened to an idle shaft 61 and being in mesh with the outputgear 13 and a second intermediate gear 63 fastened to the idle shaft 61and being in mesh with an input gear 71 of a differential gear 70. Output torque from the automatic transmission 10 is transmitted to thedifferential case 72 from the output gear 13 through these front andrear intermediate gears 61 and 63 to drive right and left axles 73 and74.

Operation of the friction coupling elements (brakes and clutches) 51-55and one-way clutch (OWCL) 56 in regard to the specified transmissiongears is described in Table I.

The transmission gear mechanism of the automatic transmission 10 shownin the skeleton diagram in FIG. 1 is practically constructed as shown inFIG. 2. As shown in FIG. 2, the automatic transmission 10 is providedwith a turbine speed sensor 305 installed in the transmission housing 11which is used in control as will be described later. This turbine speedsensor 305 at its head is installed in order to be opposite to theperiphery of a drum 51a of the forward clutch (FWCL) 51 which rotatestogether with the turbine shaft 27 in one united body and so as todetect the rotational speed of the turbine shaft 27 based on a periodicchange of the magnetic field which is generated by splines formed on theperiphery of the drum 51a.

In Table I, the low-reverse brake (LRBR) 55 is locked only for the 1stgear.

                  TABLE 1                                                         ______________________________________                                              FWCL    2-4BR    3-4CL LRBR   RVCL  OWCL                                Gear  (51)    (54)     (53)  (55)   (52)  (56)                                ______________________________________                                        1ST   ◯          (◯)                                                                            ◯                       2ND   ◯                                                                         ◯                                                   3RD   ◯    ◯                                          4TH           ◯                                                                          ◯                                          RV                           ◯                                                                        ◯                             ______________________________________                                    

Hydraulic Pressure Control Circuit

FIG. 3 is a circuit diagram showing a hydraulic pressure control systemfor supplying the working hydraulic pressure to and releasing theworking hydraulic pressure from the pressure chambers of the frictioncoupling elements 51-55 shown in FIGS. 1 and 2. It is to be noted thatamong the friction coupling elements, the 2-4 brake (2-4BR) 54, which iscomprised of a band brake, has a servo apply pressure chamber 54a and aservo release pressure chamber 54b into which the working hydraulicpressure is supplied. Specifically, when the working hydraulic pressureis supplied into only the servo apply pressure chamber 54a, the 2-4brake (2-4 BR) 54 is activated, and when the working hydraulic pressureis supplied into only the servo release pressure chamber 54b, or whenthe working hydraulic pressure is not supplied into the servo applypressure chamber 54a nor into the servo release pressure chamber 54b, aswell as when the working hydraulic pressure is supplied into both servoapply pressure chamber 54a and servo release pressure chamber 54b, the2-4 brake (2-4BR) 54 is released. Each of the remaining frictioncoupling elements 51-53 and 55 has a single pressure chamber, and islocked when the working hydraulic pressure is supplied into the pressurechamber thereof.

As shown in FIG. 3 in detail, the hydraulic control system 100 isprovided with, as the essential structural elements, a regulator valve101 for generating a specified level of line hydraulic pressure byregulating discharge hydraulic pressure of the oil pump 12, a manualshift valve 102 which is manually operated to switch the ranges, andvarious switching valves, including a low-reverse valve 103, a bypassvalve 104, a 3-4 shift valve 105 and a lockup control valve 106, forswitching the fluid paths leading to the friction coupling elements51-55, respectively, which are activated during gear shifts. Thehydraulic control system 100 is further provided with first and secondON-OFF solenoid valves (which are hereafter referred to simply as firstand second solenoid valves or SVs) 111 and 112 in order to operate theseswitching valves 103-106, a solenoid relay valve (which is hereafterabbreviated as SRV if necessary) 107 which switches the destination ofthe supply of working hydraulic pressure from the first solenoid valve111, and first, second and third duty solenoid valves (which arehereafter abbreviated as first, second and third DSVs if necessary) 121,122 and 123 which perform controlled generation, regulation anddischarge of the working hydraulic pressure to be supplied into thepressure chambers of the friction coupling elements 51-55.

The first and second solenoid valves (SVs) 111 and 112 and the first,second, and third duty solenoid valves (DSV) 121, 122 and 123 are of athree-way type which provides communication of the fluid path betweenupstream and downstream therefrom and drains the working fluid from thefluid path downstream therefrom. During draining, because the fluid pathupstream from each valve is shut off, the oil pump 12 does not dischargethe working fluid uselessly from the fluid path upstream the valve,reducing drive loss.

When each of the first and second solenoid valves (SVs) 111 and 112 isactivated or turned ON, it brings the fluid paths on upstream anddownstream sides therefrom into communication. Further, when each of thefirst, second and third duty solenoid valves (DSVs) 121, 122 and 123 isturned OFF, in other words, when the duty solenoid valve (DSV) 121, 122,123 operates at a duty rate of 0% (a rate of an ON duration of time inone ON-OFF cycle), it fully opens to bring the fluid paths on upstreamand downstream sides thereof into complete communication; when turnedON, in other words, when operates at a duty rate of 100%, it drains theworking fluid from the fluid path downstream therefrom by shutting offthe fluid path upstream thereof; and when operates at an intermediateduty rate, it generates a hydraulic pressure in the fluid pathdownstream therefrom regulated according to the duty rate by using ahydraulic pressure in the fluid path upstream therefrom as a sourcehydraulic pressure.

The line hydraulic pressure regulated through the regulator valve 101 issupplied to the manual shift valve 102 through a main hydraulic pressureline 200 as well as to a solenoid reducing valve 108 (which is hereafterreferred to simply as a reducing valve) and the 3-4 shift valve 105. Theline hydraulic pressure supplied to the reducing valve 108 is reduced toa fixed level and then supplied to the first and second solenoid valves(SVs) 111 and 112 through hydraulic pressure lines 201 and 202,respectively. While the fixed level line hydraulic pressure is suppliedto the solenoid relay valve (SRV) 107 through a hydraulic pressure line203 when the first solenoid valve (SV) 111 is ON, it is further suppliedto a control port of the bypass valve 104 as a pilot hydraulic pressurethrough a hydraulic pressure line 204 when the spool of the solenoidrelay valve (SRV) 107 is placed in its right-end position as viewed inFIG. 3 to force the spool of the bypass valve 104 toward the left-endposition. On the other hand, when the solenoid relay valve (SRV) 107places its spool in the left-end position, the fixed level linehydraulic pressure is supplied to a control port of the 3-4 shift valve105 as a pilot hydraulic pressure through a hydraulic pressure line 205to force the spool of the 3-4 shift valve 105 toward the right-endposition.

The fixed level line hydraulic pressure provided from the reducing valve108 is further supplied to the bypass valve 104 through a hydraulicpressure line 206 when the second solenoid valve (SV) 112 is ON;supplied as the pilot hydraulic pressure to a control port of the lockupcontrol valve 106 through a hydraulic pressure line 207 when the spoolof the bypass valve 104 is placed in its right-end position to force thespool of the lockup control valve 106 toward the left-end position; andsupplied to a control port of the low-reverse valve 103 through ahydraulic pressure line 208 when the spool of the bypass valve 104 isplaced its left-end positioned to force the spool of the low-reversevalve 103 toward the left-end position. Additionally, the fixed levelline hydraulic pressure from the reducing valve 108 is supplied to acontrol port 110a of the regulator valve 101 through a hydraulicpressure line 209. In this event, the fixed level line hydraulicpressure is adjusted according to, for example, opening of an enginethrottle by way of a linear solenoid valve 131 in the hydraulic pressureline 209 and accordingly, the line hydraulic pressure is adjustedaccording to throttle opening by way of the regulator valve 101.

The main hydraulic pressure line 200 leading to the 3-4 shift valve 105is brought into communication with a first accumulator 141 through ahydraulic pressure line 210 when the 3-4 shift valve 105 has placed itsspool in the right-end position to introduce the line hydraulic pressureinto the accumulator 141. On the other hand, the line hydraulic pressuresupplied to the manual shift valve 102 from the main hydraulic pressureline 200 is introduced into a first output hydraulic pressure line 211and a second output hydraulic pressure line 212 for forward ranges, i.e.the drive (D) range, the second speed (S) range and the low (L) range;into the first output hydraulic pressure line 211 and a third outputhydraulic pressure line 213 for the reverse (R) range; and into thethird output hydraulic pressure line 213 for the neutral (N) range.

The first output hydraulic pressure line 211 leads to the first dutysolenoid valve (DSV) 121 to supply the line hydraulic pressure as acontrol source hydraulic pressure to the first duty solenoid valve(IDSV) 121. The first duty solenoid (IDSV) 121 at its downstream sideleads to the low-reverse valve 103 through a hydraulic pressure line214; leads to the servo apply pressure chamber 54a of the 2-4 brake(2-4BR) 54 through a servo apply hydraulic pressure line 215 when thelow-reverse valve 103 has placed the spool in the right-end position;and further leads to the pressure chamber of the low-reverse brake(LRBR) 55 through a low-reverse brake hydraulic pressure line 216 whenthe low-reverse valve 103 has placed the spool in the left-end position.A hydraulic pressure line 217 branches off from the hydraulic pressureline 214 and leads to a second accumulator 142. The second outputhydraulic pressure line 212 leads to both second duty solenoid valve(DSV) 122 and third duty solenoid valve (DSV) 123 to supply the linehydraulic pressure as a control source hydraulic pressure to these dutysolenoid valves (DSVs) 122 and 123. The second output hydraulic pressureline 212 also leads to the 3-4 shift valve 105. The second outputhydraulic pressure line 212 leading to the 3-4 shift valve 105 isbrought into communication with the lock-up control valve 106 through ahydraulic pressure line 218 when the 3-4 shift valve 105 has placed thespool in the right-end position; and brought into communication with thepressure chamber of the forward clutch (FWCL) 51 through a forwardclutch hydraulic pressure line 219 when the lockup control valve 106 hasplaced the spool in the left-end position. A hydraulic pressure line 220branching off from the forward clutch hydraulic pressure line 219 leadsto the 3-4 shift valve 105. This hydraulic pressure line 220 is broughtinto communication with the first accumulator 141 through the hydraulicpressure line 210 when the 3-4 shift valve 105 has placed the spool inthe left-end position, and on the other hand, is brought intocommunication with the servo release pressure chamber 54b of the 2-4brake (2-4BR) 54 through a servo release hydraulic pressure line 221when the 3-4 shift valve 105 has placed the spool in the right-endposition.

The second duty solenoid valve (DSV) 122 at its downstream side to whicha control source hydraulic pressure is supplied through the secondoutput hydraulic pressure line 212 leads to a control port of thesolenoid relay valve (SRV) 107 through a hydraulic pressure line 222 andsupplies a pilot hydraulic pressure to the port to force the solenoidrelay valve (SRV) 107 to shift its spool toward the left-end position. Ahydraulic pressure line 223 branching off from the hydraulic pressureline 222 leads to the low-reverse valve 103, and is further brought intocommunication with a hydraulic pressure line 224 when the low-reversevalve 103 has placed the spool in the right-end position. A hydraulicpressure line 225 branching off from the hydraulic pressure line 224through an orifice 151 leads to the 3-4 shift valve 105 and is broughtinto communication with the servo release pressure chamber 54b of the2-4 brake (2-4BR) 54 through the servo release hydraulic pressure line221 when the 3-4 shift valve 105 has placed the spool in the left-endposition. A hydraulic pressure line 226 branching off from the hydraulicpressure line 225 leads to the bypass valve 104, and is further broughtinto communication with the pressure chamber of the 3-4 clutch (3-4CL)53 trough a 3-4 clutch hydraulic pressure line 227 when the bypass valve104 has placed the spool in the right-end position. Further, thehydraulic pressure line 224 leads directly to the bypass valve 104, andis brought into communication with the hydraulic pressure line 225through the hydraulic pressure line 226 when the bypass valve 104 hasplaced the spool in the left-end position. In other words, the hydraulicpressure lines 224 and 225 are intercommunicated with each other bybypassing the orifice 151.

The third duty solenoid valve (DSV) 123 at the downstream side to whicha control source hydraulic pressure is supplied from the second outputhydraulic pressure line 212 leads the lock-up control valve 106 througha hydraulic pressure line 228, and is brought into communication withthe forward clutch hydraulic pressure line 219 when the lock-up controlvalve 106 has placed the spool in the right-end position. On the otherhand, the third duty solenoid valve (DSV) 123 is brought intocommunication with the front pressure chamber 26a of the lock-up clutch26 through a hydraulic pressure line 229 when the lock-up control valve106 has placed the spool in the left-end position.

The third output hydraulic pressure line 213 extending from the manualshift valve 102 leads to the low-reverse valve 103 to supply the linehydraulic pressure to the low-reverse valve 103. The low-reverse valve103 directs the line hydraulic pressure to the pressure chamber of thereverse clutch (RVCL) 52 through a reverse clutch hydraulic pressureline 230. A hydraulic pressure line 231 branching off from the thirdoutput hydraulic pressure line 213 leads to the bypass valve 104, andsupplies the line hydraulic pressure as a pilot hydraulic pressure tothe control port of the low-reverse valve 103 through the hydraulicpressure line 208 when the bypass valve 104 has placed the spool in theright-end position, forcing the low-reverse valve 103 to shift the spooltoward the left-end position.

The hydraulic control circuit 100 is provided with a converter reliefvalve 109 to adjust the working hydraulic pressure supplied from theregulator 101 through a hydraulic pressure line 232 to a fixed level,and then directs the fixed level of hydraulic pressure to the lock-upcontrol valve 106 through a hydraulic pressure line 233. The fixed levelhydraulic pressure is supplied to the front pressure chamber 26a of thelock-up clutch 26 through the hydraulic pressure line 229 when the-lock-up control valve 106 has placed the spool in the right-endposition, and is supplied to the rear pressure chamber 26b through ahydraulic pressure line 234 when the lock-up control valve 106 hasplaced the spool in the left-end position. This lock-up clutch 26 isreleased when the fixed level hydraulic pressure is supplied to thefront pressure chamber 26a, and is, however, controlled to allowslippage according to the level of the working hydraulic pressuresupplied to the front pressure chamber 26a from the third duty solenoidvalve (DSV) 123 when the lock-up control valve 106 has placed the spoolin the left-end position.

A hydraulic pressure line 235, which is brought into communication withthe main hydraulic pressure line 200 through the manual valve 102 ineach of the drive (D) range, the second (S) range, the low (L) range andthe neutral (N) range, leads to a reduction port 101b of the regulatorvalve 101 to introduce the line hydraulic pressure to the reduction port101b in the respective range, so that the line hydraulic pressure isadjusted to be lower in level in these ranges than in the remainingrange, i.e. the reverse (R) range.

FIG. 4 shows the structure of a hydraulic actuator of the 2-4 brake(2-4BR) 54 in detail. As shown in FIG. 4, the hydraulic actuator has aservo-cylinder 54d and a piston 54e which is received in theservo-cylinder 54d and provided with a stem 54f secured thereto. Theservo cylinder 54b is comprised of part of the transmission housing 11and a cover member 54c fixed to the transmission housing 11 to formtherein a cylinder chamber which is divided into two sub-chambers by thepiston 54e, i.e. the servo apply pressure chamber 54a and the servorelease pressure chamber 54b. A brake band 54g, which is wrapped arounda brake-receiving member, such as a brake drum, (not shown), has one endagainst which the piston stem 54f is forced to abut and another endagainst which a fixed stem 54h fastened to the transmission housing 11abuts. A spring 54i is installed within the interior of the servorelease pressure chamber 54b to force the piston 54e toward the servoapply pressure chamber 54a so as usually to loosen the brake band 54g.The working hydraulic pressure is supplied to both or one of the servoapply pressure chamber 54a and the servo release pressure chamber 54bfrom the hydraulic control circuit 100 to tighten or loosen the brakeband 54g, locking or unlocking the 2-4 brake (2-4BR) 54. In thishydraulic actuator, especially, the piston 54e has approximately equalpressure surface areas on the sides of the servo apply pressure chamber54a and the servo release pressure chamber 54b and, therefore, when, forexample, both pressure chambers 54a and 54b are supplied with an equalworking hydraulic pressure, the piston 54e is activated only by theexpanding force of the spring 54i to move toward the servo applypressure chamber 54a, so as to loosen the band brake 54g.

FIG. 5 shows a control unit 300 provided in the automatic transmission10 which controls the first and second solenoid valves (SVs) 111 and 112and the first through third duty solenoid valves (DSVs) 121 through 123as well as the linear solenoid valve 131. The control unit 300 receivesvarious signals such as a vehicle speed signal from a speed sensor 301,a throttle opening signal from a throttle opening sensor 302, an enginespeed signal from an engine speed sensor 303, a transmission positionsignal from a shift position sensor 304, a turbine speed signal from aturbine speed sensor 305, and a temperature signal from a fluidtemperature sensor 306. With these signals, the control unit 300controls operation of each of these valves 111, 112, 121-123 and 131according to driving conditions of the vehicle or operating conditionsof the engine. Various types of these sensors are well known in the art,and any well known type may be taken. As was previously described, inparticular, the turbine speed sensor 305 may be installed as shown inFIG. 2.

The relationship between operation of these first and second solenoidvalves (SVs) 111 and 112 and the first to third duty solenoid valves(DSVs) 121 to 123 and supply of the working hydraulic pressure to eachof the friction coupling elements 51 to 55 is described in each of thepossible gears in Table II. In Table II, a mark "O" (circle) inparentheses indicates an ON state in regard to the first and secondsolenoid valves (SVs) 111 and 112 and, however, an OFF state in regardto the first to third duty solenoid valves (DSVs) 121 to 123, in eachstate, the valve bringing fluid paths upstream and downstream therefrominto communication with each other to permit a source hydraulic pressureto directly flow from the upstream path to the downstream-path. Also, amark "X" in parentheses indicates an OFF state in regard to the firstand second solenoid valves (SVs) 111 and 112 and an ON state in regardto the first to third solenoid valves (DSVs) 121 to 123, in each bothstate, the valve draining the working fluid from the upstream path whileshutting off the upstream path.

                  TABLE II                                                        ______________________________________                                        RANGE      DRIVE(Second)     LOW     REV                                      GEAR       1ST    2ND     3RD  4TH   1ST   REV                                ______________________________________                                        1ST SV (111)                                                                             X      X       X    ◯                                                                       ◯                                                                       ◯                      2ND SV (112)                                                                             X      X       X    X     ◯                                                                       ◯                      1ST DSV (121)                                                                            X      ◯                                                                         ◯                                                                      ◯                                                                       ◯                                                                       ◯                      2ND DSV (122)                                                                            X      X       ◯                                                                      ◯                                                                       X     ◯                      3RD DSV (123)                                                                            ◯                                                                        ◯                                                                         ◯                                                                      X     ◯                                                                       ◯                      ______________________________________                                    

As shown in FIG. 6 and indicated in Table II, for the first (1st) gearin the forward ranges excepting the low (L) range, only the third dutysolenoid valve (DSV) 123 operates to generate a working hydraulicpressure from the line hydraulic pressure as a source hydraulic pressurefrom the second output hydraulic pressure line 212. This workinghydraulic pressure is supplied to the lock-up control valve 106 throughthe hydraulic pressure line 228. Because, at this time, the lock-upcontrol valve 106 has placed the spool in the right-end position, theworking hydraulic pressure is directed to the pressure chamber of theforward clutch (FWCL) 51 as a forward clutch hydraulic pressure throughthe forward clutch hydraulic pressure line 219, locking the forwardclutch (FWCL) 51. In this instance, because the hydraulic pressure line220 branching off from the forward clutch hydraulic pressure line 219has been brought into communication with the first accumulator 141through the hydraulic pressure line 210 via the 3-4 shift valve 105, theforward clutch hydraulic pressure is supplied smoothly.

For the second (2nd) gear, as shown FIG. 7 and indicated in Table II, inaddition to the third duty solenoid valve (DSV) 123 locked in the first(1st) gear, the first duty solenoid valve (DSV) 121 operates to generatea working hydraulic pressure from the line hydraulic pressure as asource hydraulic pressure from the first output hydraulic pressure line211. This working hydraulic pressure is supplied to the low-reversevalve 103, and at this time, because the low-reverse valve 103 hasplaced the spool in the right-end position, the working hydraulicpressure is directed to the servo apply hydraulic pressure line 215, andthen supplied to the servo apply pressure chamber 54a of the 2-4 brake(2-4BR) 54 and locking the 2-4 brake (2-4BR) 54, while the forwardclutch (FWCL) 51 is locked. In this instance, because the hydraulicpressure line 214 leads to the second accumulator 142 through thehydraulic pressure line 217, it is gentle to supply the servo applyhydraulic pressure line 215, and hence to lock the 2-4 brake (2-4BR) 54.The working fluid accumulated in the accumulator 142 is pre-charged tothe pressure chamber of the low-reverse brake (LRBR) 55 through thelow-reverse brake hydraulic pressure line 216 when the low-reverse valve103 shifts the spool toward the left-end position during a gear shift tothe first (1st) gear in the low (L) range, as will be described later.

For the third (3rd) gear, as shown FIG. 8 and indicated in Table II,while the first and second solenoid valves (SVs) 111 and 112 and thefirst and third duty solenoid valves (DSVs) 121 and 123 remains in thestate of the second (2nd) gear, the second duty solenoid valve (DSV) 122operates to generate a working hydraulic pressure from the linehydraulic pressure as a source hydraulic pressure-supplied from thesecond output hydraulic pressure line 212. This working hydraulicpressure is supplied to the low-reverse valve 103 through the hydraulicpressure lines 222 and 223, and then, because the low-reverse valve 103still remains the spool in the right-end position, directed to thehydraulic pressure line 224. As a result, the working hydraulic pressureis introduced into the hydraulic pressure line 225 through the orifice151 from the hydraulic pressure line 224, and then to the 3-4 shiftvalve 105. At this time, because the 3-4 shift valve 105 has placed- thespool in the left-end position, the working hydraulic pressure isfurther directed as a servo release hydraulic pressure to the servorelease pressure chamber 54b of the 2-4 brake (2-4BR) 54 through theservo release hydraulic pressure line 221. Consequently, the 2-4 brake(2-4BR) 54 is unlocked or released. On the other hand, the workinghydraulic pressure is directed to the bypass valve 104 through thehydraulic pressure line 226 branching off from the hydraulic pressureline 225 after the orifice 151. At this time, because the bypass valve104 has been placed the spool in the right-end position, the workinghydraulic pressure is also supplied as a 3-4 clutch hydraulic pressureto the pressure chamber of the 3-4 clutch (3-4CL) 53 through the 3-4clutch hydraulic pressure line 227. In this way, while the 2-4 brake(2-4BR) 54 is unlocked, both forward clutch (FWCL) 51 and 3-4 clutch(3-4CL) 53 are locked. In this instance, in the sate of operation of thevalves for the third (3rd) gear, the second duty solenoid valve (DSV)122 generates the working hydraulic pressure, as was previouslydescribed, and supplies it to the relay valve 107 at the control port107a through the hydraulic pressure line 222 to force the relay valve107 to shift the spool to the left-end position.

For the forth (4th) gear, as shown in FIG. 9 and indicated in Table II,while the valves 112, 121 and 122 remain in the same state of operationas for the third (3rd) gear, the third duty solenoid valve (DSV) 123stops the generation of working hydraulic pressure, and, on the otherhand, the first solenoid valve (SV) 111 operates. Consequently, thefirst solenoid valve (SV) 111 supplies the fixed level hydraulicpressure to the relay valve 107 from the hydraulic pressure line 201through the hydraulic pressure line 203. At this time, because the relayvalve 107 has placed the spool in the left-end position for the third(3rd) gear, it directs the fixed level hydraulic pressure to the 3-4shift valve 105 at the control port 105a through the hydraulic pressureline 205, forcing the 3-4 shift valve 105 to shift the spool to theright-end position, so that the servo release hydraulic pressure line221 is brought into communication with the hydraulic pressure line 220branching off from the forward clutch hydraulic pressure line 219 to putthe releasing chamber 54b of the 2-4 brake (2-4BR) 54 and the pressurechamber of the forward clutch (FWCL) 51 intercommunicated with eachother.

By means of putting the third duty solenoid valve (DSV) 123 inoperativeto stop the generation of working hydraulic pressure and drain theworking fluid from the downstream path from the third duty solenoidvalve (DSV) 123, while the servo release hydraulic pressure is drainedfrom the releasing chamber 54b of the 2-4 brake (2-4BR) 54 through thehydraulic pressure line 228 via the lock-up control valve 106, to lockthe 2-4 brake (2-4BR) 54 again, and the forward clutch hydraulicpressure is drained from the pressure chamber of the forward clutch(FWCL) 51 through the hydraulic pressure line 228 via the lock-upcontrol valve 106 to unlock the forward clutch (FWCL) 51.

For the first (1st) gear in the low (L) range, as shown in FIG. 10 andindicated in Table II, the first and second solenoid valves (SVs) 111and 112, and the first and third duty solenoid valves (DSVs) 121 and 123operate to supply the working hydraulic pressure generated by way of thethird duty solenoid valve (DSV) 123 as a forward clutch hydraulicpressure to the pressure chamber of the forward clutch (FWCL) 51 throughthe hydraulic pressure line 228 and forward clutch hydraulic pressureline 219 via the lock-up-control valve 106 in a similar way for thefirst (1st) gear, for example, in the drive (D) range. In this manner,the forward clutch (FWCL) 51 is applied with the working hydraulicpressure to lock. At this time, due to accumulation of the workinghydraulic pressure in the first accumulator 141 through the hydraulicpressure lines 210 and 220 via the 3-4 shift valve, the forward clutch(FWCL) 51 is locked smoothly.

By means of the operation of the first solenoid valve (SV) 111, thebypass valve 104 at the control port 104a is supplied with a pilothydraulic pressure through the-hydraulic pressure lines 203 and 204 viathe relay valve 107 to shift the spool to the left-end position, whichis followed by introduction of the working hydraulic pressure into thehydraulic pressure line 208 through the hydraulic pressure line 206 viathe bypass valve 104, forcing the low reverse valve 103 to shift thespool to the left-end position. Consequently, the working hydraulicpressure at the first duty solenoid valve (DSV) 121 is supplied as alow-reverse brake hydraulic pressure to the pressure chamber of thelow-reverse brake (LRBR) 55 through the low-reverse brake hydraulicpressure line 216 the low-reverse valve 103, by which, while the forwardclutch (FWCL) 51 is locked, the low-reverse brake (LRBR) 55 is locked,providing the first (1st) gear where engine brake is available.

In the reverse (R) range, as shown in FIG. 11 and indicated in Table II,all of the valves 111 and 112, and 121 to 123 operate. In this state,the second and the third duty solenoid valves (DSVs) 122 and 123 do notgenerate any working hydraulic pressure due to interruption of supply ofthe source hydraulic pressure thereto from the second output hydraulicpressure line 212. As was described, because the first and secondsolenoid valves (SVs) 111 and 112 operate, the bypass valve 104 shiftsthe spool to the left-end position similarly for the first (1st) gear inthe low (L) range, which is followed by forcing the low reverse valve103 to shift the spool to the left-end position. Under thiscircumstance, the working hydraulic pressure generated at the first dutysolenoid valve (DSV) 121 is supplied as the low-reverse brake hydraulicpressure to the pressure chamber of the low-reverse brake (LRBR) 55. Inthe reverse (R) range, the line hydraulic pressure is introduced intothe third output hydraulic pressure line 213 from the manual shift valve102, and is directed as a reverse clutch hydraulic pressure to thepressure chamber of the reverse clutch (RVCL) 53 through the reverseclutch hydraulic pressure line 230 via the low-reverse valve 103 withthe spool shifted to the left-end position. Consequently, the reverseclutch (RVCL) 52 and the low reverse brake (LRBR) 55 are locked. In thisinstance, the line hydraulic pressure is introduced into the thirdoutput hydraulic pressure line 213 from the manual valve 102 even in theneutral (N) range, locking the reverse clutch (RVCL) 52 in the neutral(N) range when the low reverse valve 103 has placed the spool to theleft-end position.

Control Operation

The following description is directed to gear shifts, in particular upshifts, through the gear shift control accomplished by the control unit300. Specifically, as shown in FIG. 12, up shift control is accomplishedby feedback controlling mainly supply of the working pressure to thefriction coupling element to be locked so as to bring a change rate dNtof the dropping turbine speed Nt into agreement with a target changerate dNt₀ as shown in FIG. 12. The turbine speed change rate dNt isequivalent to a change in transmission output torque T₀ (which ishereafter refereed to simply as an output torque change) .increment.T₀between during an inertia phase for which the turbine speed Nt changesdue to a gear shift and after a conclusion of the gear shift as shown inFIG. 13. If the output torque change .increment.T₀ becomes larger thanthe output torque difference between before and after the gear shift, ashift shock is enhanced. On the other hand, if it is smaller than thedifference, a shift time necessary to conclude the gear shift isprolonged. Accordingly, the target turbine speed change rate dNt₀ is setto meet the output torque change .increment.T₀ so as to be approximatelyequal to the output torque difference in transmission output torquebetween before and after the gear shift.

(1) 1-2 Up Shift

As apparent from FIGS. 6 and 7, a 1-2 gear shift is performed by, whilelocking the forward clutch 51 with a working pressure generated throughthe third duty solenoid valve 123, supplying a servo apply pressuregenerated through the first duty solenoid valve 121 to the servo applypressure chamber 54a of the 2-4 brake 54. In this instance, the servoapply pressure is feedback controlled through the first duty solenoidvalve 121. As was previously mentioned, when each of the first, secondand third duty solenoid valves 121, 122 and 123 operates at a duty rateof 0%, it fully opens to bring the fluid paths on upstream anddownstream sides thereof into complete communication; when operates at aduty rate of 100%, it drains the working fluid from the fluid pathdownstream therefrom by shutting off the fluid path upstream thereof;and when operates at an intermediate duty rate, it generates a hydraulicpressure in the fluid path downstream therefrom regulated according tothe duty rate by using a hydraulic pressure in the fluid path upstreamtherefrom as a source hydraulic pressure.

The feedback control of servo apply hydraulic pressure by way of thefirst duty solenoid valve 121 is performed following the flowchartillustrating a sequence routine shown in FIG. 14. Explaining thissequence routine with reference to a time chart shown in FIG. 15, when a1-2 gear shift command is given, a calculated hydraulic pressure Ps isderived by adding a base hydraulic pressure Pb, a feedback hydraulicpressure Pfb and a learned pressure Pad all together at step S3subsequently to calculations of the base hydraulic pressure Pb, thefeedback hydraulic pressure Pfb and made at steps S1, S2 and S3,respectively. The calculations of these base hydraulic pressure Pb,feedback hydraulic pressure Pfb and learned pressure Pad will bedescribed in connection with hydraulic pressure calculation subroutinesshown in FIGS. 16, 20 and 25, respectively. As shown in FIG. 15, thebase pressure Pb is kept at a constant initial level of pressure Pb'from an appearance of the gear shift command to a point of time at whichthe turbine speed starts to fall, i.e. at which an inertia phase startsas labeled "A" in FIG. 15 and increases at a fixed rate after thecommencement of the inertia phase. The feedback pressure Pfb isestablished so as to bring the turbine speed change rate dNt intoagreement with the target turbine speed change rate dNt₀ during theinertia phase after a passage of a specified time interval T1 from thecommencement of the inertia phase. The reason for not performing thecalculation of the feedback pressure Pfb until passage of the specifiedtime interval T1 from the commencement of the inertia phase is that itis impossible to obtain a precise turbine speed change rate dNt, whichis-used in the calculation as a foundation of the base pressure, at thebeginning of the inertia phase. Further, the learned pressure Pad is seton the basis of the state of an inertia phase in a first previous 1-2gear shift after a conclusion of the first previous 1-2 gear shift. Allof these calculations will be described in detail later.

Subsequently, a determination is made at step S5 whether or not aprecharge flag Fp has been up or set to a state of "1," this indicatesthat it is during the precharge control since an appearance of a gearshift command. The precharge control is performed by filling the fluidpath leading to the servo apply pressure chamber 54a of the 2-4 brake 54with a working fluid immediately after the beginning of a gear shift inorder to provide an improved gear shift responsiveness. If the prechargeflag Fp is up, the first duty solenoid valve 121 is operated at a dutyrate of 0% to remain fully open at step S6. On the other hand, when theprecharge flag Fp is down or has been reset to a state of "0," thisindicates that the precharge control has been concluded, then, anotherdetermination is made at step S7 whether or not the 1-2 gear shift hasbeen concluded. The criterion for the determination of up-shiftconclusion may be achievement of any one of such events that the turbinespeed change rate dNt has turned over from negative to positive, thatthe absolute value of the turbine speed change rate dNt has fallen lowerthan half a turbine speed change rate dNt during the gear shift is inprogress, and that the turbine speed Nt has fallen below a turbine speed(which is hereafter referred to as a shift-end turbine speed) at the endof gear shift which is calculated based on the turbine speed at thebeginning of the gear shift. When the 1-2 gear shift has not yet beenconcluded after the conclusion of the precharge control, the first dutysolenoid valve 121 is operated at a duty rate corresponding to thecalculated hydraulic pressure Ps at step S8. On the other hand, after aconclusion of the 1-2 gear shift, the duty rate is reduced at a fixedrate and returned once again to 0% through steps S9 and S10. Tn thismanner, the servo apply hydraulic pressure is controlled as shown inFIG. 15 to bring the turbine speed change rate dNt during the inertiaphase into agreement with a target turbine speed change rate dNt₀.

(2) Base Hydraulic pressure Calculation

The calculation of the base hydraulic pressure Pb is accomplishedfollowing the flowchart illustrating the base hydraulic pressurecalculation sequence subroutine shown in FIG. 16.

At step S11, a target turbine speed change rate dNt₀ during the gearshift is calculated, and then, at step S12, a hydraulic pressure Pi forthe target turbine speed change rate dNt₀ is read from a hydraulicpressure map such as shown in FIG. 17 in which the greater the hydraulicpressure Pi is, the greater the absolute value of the target turbinespeed change rate dNt₀ is. Subsequently, hydraulic pressures Pt and Pt2corresponding to target turbine torque Tt₀ during the gear shift and itssquare (Tt₀)² are read from hydraulic pressure maps such as shown inFIGS. 18 and 19, respectively, at steps S13 and S14, respectively. Aninitial base hydraulic pressure Pb' is found by adding these hydraulicpressures Pt and Pt2 to the hydraulic pressure Pi at step S15. In thisinstance, this target turbine torque Tt₀ is calculated by multiplyingthe turbine torque before a gear shift by an engine torque drop rate.Through the correction of the hydraulic pressure Pi for the targetturbine speed change rate dNt₀ making use of the hydraulic pressures Ptand Pt2 corresponding to this target turbine torque Tt₀, fluctuations oftransmission output torque is more effectively suppressed during thegear shift.

Thereafter, at step S16, a determination is made in order to find apoint of time (labeled "A" in FIG. 15) at which the turbine speed Ntstarts to fall due to commencement of the inertia phase as to whether ornot the target turbine speed change rate dNt₀ is less than a specifiedrate K1. Until the target turbine speed change rate dNt₀ becomes lessthan the specified rate K1, the initial base pressure Pb' is directlyused as the base hydraulic pressure Pb at step S17. On the other hand,when the target turbine speed change rate dNt₀ is already less than thespecified rate K1, then, the base hydraulic pressure Pb is increased ata fixed rate by adding to the initial base pressure Pb' a correctionpressure (K2×t) obtained by multiplying a specified value K2 by the atime t passed from the point of time at which the target turbine speedchange rate dNt₀ exceeded the specified rate K1. As a result, the basepressure Pb is controlled to change as shown in FIG. 15.

(3) Feedback Hydraulic Pressure Calculation

The calculation of the feedback hydraulic pressure Pfb is accomplishedfollowing the flowchart illustrating the feedback hydraulic pressurecalculation sequential subroutine shown in FIG. 20. The feedbackpressure Pfb is calculated in a fuzzy control method in order to copewith the fact that the feedback control system has dynamiccharacteristics different in accordance with driving conditionsincluding the turbine speed Nt and so forth.

The flowchart logic commences and control passes directly to a functionblock S21 where a calculation of the deviation Ed(t) of the turbinespeed change rate dNt from the target turbine speed change rate dNt₀. Atstep S22, fuzzy stratification is made on turbine speeds Nt to classifythem into a plurality of sheaves according to which three functions ofmembers Mn1, Mn2 and Mn3 have different grade values between one andzero. These membership functions Mn1, Mn2 and Mn3 are graphically shownin FIG. 21 and defined by the following numerical formulas: ##EQU1##

Subsequently, at step S23, feedback-manipulated variables Fb1(t), Fb2(t)and Fb3(t) in accordance with the deviations Ed(t) obtained at step T201by calculating the following generalized expressions F1, F2 and F3 inwhich coefficients corresponding to the respective three regions ofturbine speeds Nt having different dynamic characteristics are given.

    Fb1(t)=F1[Ed(t), Ed(t-1), Ed(t-2), . . . Fb1(t-1), Fb1(t-2), Fb1(t-3), . . . ]

    Fb2(t)=F2[Ed(t), Ed(t-1), Ed(t-2), . . . Fb2(t-1), Fb2(t-2), Fb2(t-3), . . . ]

    Fb3(t)=F3[Ed(t), Ed(t-1), Ed(t-2), . . . Fb3(t-1), Fb3(t-2), Fb3(t-3), . . . ]

These generalized expressions are used to calculate the latestfeedback-manipulated variables Fb1(t), Fb2(t) and Fb3(t) by substitutingthe latest and previous deviations Ed(t) and Ed(t-i) and the previousfeedback-manipulated variables Fb1(t-i), Fb2(t-i) and Fb3(t-i) into therespective functions.

Finally, at step S24, eventual feedback-manipulated variables Fb(t) iscalculated by performing fuzzy composition of the latestfeedback-manipulated variables Fb1(t), Fb2(t) and Fb3(t) in accordancewith the following formula in which grade values Mn1(Nt), Mn2(Nt) andMn3(Nt) of the respective membership functions Mn1, Mn2 and Mn3 for thelatest turbine speeds Nt.

    Fb(t)=[Fb1(t)·Mn1(Nt)+Fb2(t)·Mn2(Nt)+Fb3(t)·Mn3(Nt)]÷[Mn1(Nt)+Mn2(Nt)+Mn3(Nt)]

This eventual feedback-manipulated variable Fb(t) is substituted for thefeedback pressure Pfb. As apparent from the above description, becausethe feedback pressure Pfb is calculated by composing the latestfeedback-manipulated variables Fb1(t), Fb2(t) and Fb3(t) which have beenweighted in accordance with the regions of turbine speeds Nt, even inthe event that dynamic characteristics of the feedback control system,such as a characteristic of turbine speed changing with respect to theworking pressure and a characteristic of the working pressure withrespect to the duty rate of a pressure control duty solenoid valve forexample, are different in accordance with the regions of turbine speedsNt, the feedback control is accomplished under the dynamiccharacteristics peculiar to a specific region of turbine speeds.

The feedback-manipulated variables Fb1(t), Fb2(t) and Fb3(t) arepractically calculated from the following formulas: ##EQU2## Theseformulas, which are expressed in the form of what is called a transferfunction, include coefficients A1₀ through A1₆ and B1₀ through B1₆, A2₀through A2₆ and B2₀ through B2₆, and A3₀ through A3₆ and B3₀ throughB3₆, respectively, each of the groups of coefficients being establisheddifferently so as to correspond to dynamic characteristics for therespective turbine speed regions. The feedback-manipulated variablesFb1(t), Fb2(t) and Fb3(t) may be alternatively calculated from theformulas expressed in a form of integral-proportional differential(I-PD) control as follows: ##EQU3## The groups of coefficients A1₀through A1₆ and B1₀ through B1₆, A2₀ through A2₆ and B2₀ through B2₆,and A3₀ through A3₆ and B3₀ through B3₆ are differently established soas to correspond to dynamic characteristics for the respective turbinespeed regions.

The following description is directed to the example of calculating aneventual feedback-manipulated variable Fb(t) in the fuzzy composingmanner by use of formulas corresponding to sheaves of turbine speeds Ntand sheaves of turbine torque Tt to which driving conditions are fuzzystratified in accordance with dynamic characteristics of the feedbackcontrol system. In this example, membership functions Mn1 and Mn2, eachof which includes a turbine speed Nt as a parameter, and membershipfunctions Mt1 and Mt2, each of which includes turbine torque Tt as aparameter, are graphically shown in FIGS. 22 and 23 and defined by thefollowing numerical formulas: ##EQU4##

Feedback-manipulated variables Fb11(t), Fb12(t), Fb21(t) and Fb22(t) areobtained by calculating the following formulas established for the fourregions Z11, Z12, Z21 and Z22 stratified using as parameters turbinespeeds Nt and turbine torque Tt such as shown in FIG. 24 in which thedeviation Ed(t) of the turbine speed change rate from the target turbinespeed change rate and the turbine speed change rate dNt(t) aresubstituted. ##EQU5## The groups of coefficients D1₁ through D1₃, D2₁through D2₃, D3₁ through D3₃ and D4₁ through D4₃ are differentlyestablished so as to provide the feedback-manipulated variables Fb11(t),Fb12(t) and Fb21(t) and Fb22(t) correspondingly to dynamiccharacteristics for the respective turbine speed regions.

Eventual feedback-manipulated variable Fb(t) is calculated by performingfuzzy composition of the latest feedback-manipulated variables Fb11(t),Fb12(t), and Fb21(t) and Fb22(t) in accordance with the followingformula in which grade values Mn1(Nt) and Mn2(Nt) for the latest turbinespeeds Nt and Mt1(Tt) and Mt2(Tt) for the latest turbine torque Tt ofthe respective membership functions Mn1, Mn2, Mt1 and Mt2. ##EQU6## Inaccordance with to this example, because the eventualfeedback-manipulated variable Fb(t) is established in consideration withthe difference among dynamic characteristics in accordance with theregions of turbine speeds and turbine torque, the feedback control ofservo apply pressure is accomplished in conformity with drivingconditions in any region of turbine speed and turbine torque.

(4) Learned Hydraulic Pressure Calculation

At the beginning of the 1-2 up shift, the torque phase is actualized bymeans of supplying the initial base hydraulic pressure Pb' to the servoapply pressure chamber of the 2-4 brake 54 after a conclusion of theprecharge control. Practically, as shown at steps S3 and S4 of theflowchart in FIG. 14, the hydraulic pressure to be supplied is the basehydraulic pressure Pb added by the learned hydraulic pressure Padobtained during the first previous 1-2 up shift.

The learned hydraulic pressure Pad is established at conclusion of afirst previous up shift in accordance with how the feedback control wasperformed in the inertia phase of the first previous up shift.Calculation of the learned hydraulic pressure Pad is accomplishedfollowing the flowchart illustrating the sequence subroutine shown inFIG. 25. The flowchart logic commences and control passes directly to afunction block S31 a determination is made as to whether or not a gearshift has been concluded. When the gear shift has been concluded, atstep S32, detection is made to find a first peak value (absolute value)dNtp of turbine speed change rate dNt in the inertia phase and thenumber of control cycles counted before the first peak value dNtpappears in an inertia phase such as graphically shown in FIG. 26.Subsequently, membership functions Mp1 and Mp2, each of which includesthe point of time of the appearance of peak value as a parameter, areestablished in accordance with the following formulas at step T303.##EQU7## The membership functions Mp1 and Mp2 are graphicallyillustrated in FIG. 27.

At step S34, in accordance with the number of control cycles Tp,correction values Pxp1 and Pxp2 are obtained by calculating thefollowing formulas into which the peak value dNtp is substituted.

    Pxp1=-(dNtp-dNt.sub.0)·E1

    Pxp2=-(dNtp-dNt.sub.0)·E2

where E1 and E2 are coefficients differently established in accordancewith the number of control cycles Tp until an appearance of first peakvalue dNtp.

Thereafter, after striking the average Mfb of feedback-manipulatedvariables (feedback hydraulic pressures Pfb) in the inertia phase atstep S35, membership functions Mf1, Mf2 and Mf3 including the averageMfb as a parameter are established at step S36 as follows: ##EQU8##These membership functions Mf1, Mf2 and Mf3 are graphically illustratedin FIG. 28.

At step S37, correction values Pxf1, Pxf2 and Pxf3 in accordance withthe average Mfb of feedback-manipulated variables are obtained bycalculating the following formulas into which average Mfb offeedback-manipulated variables is substituted.

    Pxf1=-Mfb·F1

    Pxf2=-Mfb·F2

    Pxf3=0

where F1 and F2 are coefficients differently established in accordancewith the average Mfb of feedback-manipulated variables, and F3 takes 0(zero) regardless of the average Mfb of feedback-manipulated variables.

Following striking the average Med of deviations Ed of the turbine speedchange rates dNt with respect to the target turbine speed change ratesdNt₀ in the inertia phase at step S38, membership functions Mf1, Mf2 andMf3 including the average Med of deviations Ed as a parameter areestablished at step S39 as follows: ##EQU9## These membership functionsMe1, Me2 and Me3 are graphically illustrated in FIG. 29.

At step S40, correction values Pxe1, Pxe2 and Pxe3 in accordance withthe average Med of deviations Ed are obtained by calculating thefollowing formulas into which the average Med of deviations Ed issubstituted.

    Pxe1=Med·dNt.sub.0 ·G1

    Pxe2=-Med·dNt.sub.0 ·G2

    Pxe3=0

where G1 and G2 are coefficients differently established in accordancewith the average Med of deviations Ed, and F3 takes 0 (zero) regardlessof the average Med of deviations Ed.

Finally, at step S41, eventual correction value Pad is calculated byperforming fuzzy composition of the correction values Pxp1, Pxp2, Pxf1through Pxf3, and Pxe1 through Pxe3 in accordance with the followingformula: ##EQU10## This learned hydraulic pressure Pad is added as aneventual correction pressure to the base pressure Pb during the sametype of next gear shift to establish the working pressure during thetorque phase. In this case, because the learned hydraulic pressure Padis obtained by means of the fuzzy composition of correction valuescalculated based on a peak value dNtp of turbine speed change rate dNthaving occurred first in the inertia phase, the number of control cyclesTp until an appearance of the first peak value dNtp, the average Mfb offeedback-manipulated variables (feedback hydraulic pressure Pfb) duringthe inertia phase, the average Med of deviations Ed of the turbine speedchange rate dNt from the target turbine speed change rate dNt₀ duringthe inertia phase, respectively, during an inertia phase in the sametype of next gear shift, the feedback control is satisfactorilyaccomplished by supplying as a working pressure the base hydraulicpressure Pb added by the learned hydraulic pressure Pad, making itpossible to reduce the peak value dNtp of turbine speed change rate dNtand/or to reduce the feedback hydraulic pressure Pfb or the deviation Edof the turbine speed change rate dNt from the target turbine speedchange rate dNt₀. Consequently, precise agreement is yielded between theturbine speed change rate dNt and the target turbine speed change ratedNt₀ during the inertia phase. Although the learned hydraulic pressurePad is calculated for each type of gear shift and used in the same typeof next gear shift, it may be done to classify learned hydraulicpressures Pad in accordance with magnitude of turbine torque for thesame type of gear shifts and store them and used afterward for the sameturbine torque Tt in the same type of next gear shift. This makes itmore precisely to perform not only the working pressure control during atorque phase but also the feedback pressure control during an inertiaphase. The calculations of these base pressure Pb, feedback pressure Pfband learned pressure Pad are performed in other gear shifts as well asin the 1-2 up shift.

(5) Working Pressure Correction at the Beginning of Inertia Phase

During an inertia phase in a 1-2 up shift, while keeping the turbinespeed change rate dNt in agreement with a target turbine speed changerate dNt₀, the turbine speed Nt is reduced to a shift-end turbine speedby feedback controlling the working hydraulic pressure usingfeedback-manipulated variables. However, there are cases where, because,for example, the working pressure during a torque phase, i.e. theinitial base pressure Pb' (see FIG. 15), is improper, the feedbackcontrol is unsatisfactorily performed after a shift to an inertia phasefrom the torque phase. For example, if the initial base hydraulicpressure Pb' is too high, the working hydraulic pressure is at a highlevel at the beginning of an inertia phase due to a delay of control insuch a direction as to reduce the working hydraulic pressure at thebeginning of the feedback control, causing the 2-4 brake 54 to bequickly. To the contrary, if the initial base pressure Pb' is too low,the inertia phase expends a long time due to a delay of control in sucha direction as to increase the working hydraulic pressure at thebeginning of the feedback control. In either case, it is hard to providea satisfactory shift feeling.

In view of the above circumstances, in this embodiment, the workingpressure correction control is performed in accordance with states ofthe feedback control at the beginning of an inertia phase to accomplishsatisfactorily the feedback control during the inertia phase afterdetection of the state of the feedback control.

The working pressure correction at the beginning of an inertia phase isaccomplished following the flowchart illustrating the sequence routineshown in FIG. 30. This working pressure correction is performed alongwith the servo apply pressure feedback control shown in FIG. 14 for a1-2 up shift to correct the feedback hydraulic pressure Pfb calculatedat step S2 of the flowchart. The flowchart logic commences in responseto an appearance of a 1-2 gear shift command and control passes directlyto a function block at S51 where initialization is made to clear theintegrated value Sdnt of turbine speed change rate Nt and the integratedvalue Sdnt₀ of target turbine speed change rate dNt₀. Thereafter, adetermination is made at step S52 as to whether or not a torque phasehas been over. This determination is practiced by detecting a point oftime at which the turbine speed change rate dNt reverses from positiveto negative (see a point labeled "A" in FIG. 31). When the feedbackcontrol has shifted to an inertia phase from the torque phase, throughsteps S53 to S55, integration is performed on the turbine speed changerate dNt and the target turbine speed change rate dNt₀ to calculatetheir integration values Sdnt (=Sdnt+dNt) and Sdnt₀ (=Sdnt₀ +dNt₀) untilpassage of a specified time T2 (which is, for example, 75 ms when thecorrection control needs 25 ms to perform one control cycle) from thephase shift. After the passage of the specified time T2, at step S56,deviation Es of the integration value Sdnt of the turbine speed changerate dNt with respect to the integration value Sdnt₀ of target turbinespeed change rate dNt₀. Subsequently, a determination is made at stepS57 as to whether or not the deviation Es has an absolute value greaterthan a specified value K3. The absolute value of deviation Es isequivalent to an area shadowed in FIG. 31.

When the absolute value of deviation Es is greater than the specifiedvalue K3, this indicates that there is great deviation between theturbine speed change rate dNt and the target turbine speed change ratedNt₀ and it is regarded that the feedback control is possiblyunsatisfactorily performed afterward, then, after finding a correctionpressure Px in accordance with the deviation Es by calculating anequation Px=Es×K4 (K4: constant) at step S58, the feedback hydraulicpressure Pfb is corrected by adding the correction pressure Px thereto.

The case as shown in FIG. 31 where the deviation Es between theintegration values Sdnt and Sdnt₀ has a minus value occurs when theworking pressure is too high. In this case, as indicated bydotted-broken line "B" in FIG. 31, the absolute value of turbine speedchange rate dNt becomes greater than the absolute value of targetturbine speed change rate dNt₀, producing a rapid drop in turbine speedNt. However, because the feedback pressure Pfb is added by a minuscorrection pressure Px, it is momentarily forced to drop as labeled "C"in FIG. 31. As a result, when the feedback control is resumed, theturbine speed change rate dNt rapidly converges at the target turbinespeed change rate dNt₀. In a similar manner, even when the deviation Eshas a minus value, the turbine speed change rate dNt rapidly convergesat the target turbine speed change rate dNt₀. The working pressurecorrection is performed at the beginning of inertia phase during upshifts as well as during the 1-2 up shift.

(6) Setting of Precharge Period

The precharge period correction performed coincidentally with the servoapply pressure feedback control shown in FIG. 14 at the beginning ofgear shift is accomplished following the flowchart illustrating thesequence routine shown in FIG. 32. The flowchart logic commences inresponse to an appearance of a gear shift command and control passesdirectly to a function block at S61 where initialization is made toreset the total flowing quantity of working fluid Qt. Subsequently, atstep S62, a base flowing quantity Q, which is defined as the flowingquantity of working fluid passing through the first duty solenoid valve121 operating at a duty rate of 0%, in accordance with the line pressureis read from a flowing rate map shown in FIG. 33. In the flowing ratemap, the base flowing quantity of working fluid is specified to becomegreater with an increase in line pressure. This is because, even whenthe first duty solenoid valve 121 operates at a duty rate of 0%, theflowing quantity of working fluid passing through the first dutysolenoid valve 121 changes in accordance with the level of line pressureand increases with an increase in line pressure. At step S63, a flowingrate correction coefficient K5 is read from a flowing rate correctionmap shown in FIG. 34 in order to correct the base flowing rate inaccordance with the temperature of working fluid. The flowing ratecorrection map specifies the flowing rate correction coefficient K5 soas to lower it gradually from 1 (one) as the temperature of workingfluid drops. After calculating a corrective flowing quantity Qx bymultiplying the base flowing quantity Q by the correction coefficient K5at step S64, a total flowing quantity Qt is obtained by calculating thefollowing formula at step S65.

    Qt=Qt(t-1)+Qx

where Qt(t-1) is the total flowing quantity in the previous cycle ofsequential routine.

In cases where the working fluid passes through a valve at a flowingrate lower when it has a low viscosity due to a low temperature than ithas a standard viscosity, the base flowing quantity is corrected todecrease in conformity with actual circumstances, so as to be adjustedto a practical flowing quantity.

Thereafter, a determination is made at step S66 as to whether or not thetotal flowing quantity Qt has exceeded a specified flowing quantity K6.Until the total flowing quantity Qt exceeds the specified flowingquantity K6, the precharge flag Fp remains up at step S67. On the otherhand, when the total flowing quantity Qt exceeds the specified flowingquantity K6, the precharge flag Fp is reset to the state of "0" at stepS68. In this instance, the specified flowing quantity K6 is establishedsuch as to be equal to the volume of fluid in a fluid path between thespecific valve of the hydraulic pressure control circuit 100 to apressure chamber of the specific friction coupling element, namely thehydraulic pressure line from the first duty solenoid valve 121 to theservo apply pressure chamber 54a of the 2-4 brake 54, when the gearshift is a 1-2 up shift. Consequently, it is regarded that the fluidpath is filled with the working fluid when the total flowing quantity Qtreaches the specified flowing quantity K6, the precharge flag Fp isreset to the state of "0" at the moment.

In the period that the precharge flag Fp remains up, the first dutysolenoid valve 121 is controlled to operate at a duty rate of 0% at stepS6 of the flowchart shown in FIG. 14, quickly filling the hydraulicpressure line leading to the servo apply pressure chamber 54a of the 2-4brake 54 with the working fluid. The precharge period setting routine isexecuted if necessary during up shifts as well as during the 1-2 upshift.

(7) 2-3 Up Shift

Basically, during a 2-3 up shift, the hydraulic pressure control circuit100, on one hand, locks the 3-4 clutch 53 and unlocks the 2-4 brake 54with the remaining friction coupling elements remained in the state forthe second (2nd) gear. In order to lock and unlock the 3-4 clutch 53 andthe 24 brake 54, respectively, the second duty solenoid valve 122 iscontrolled to generate hydraulic pressures necessary to lock and unlockthese 3-4 clutch 53 and 2-4 brake 54, respectively. During the 2-3 upshift, the height of what is called a shelf pressure, which refers tothe fact that the hydraulic pressure in an inertia phase during lockingthe 3-4 clutch 53 increasingly or decreasingly changes through anintermediate shift constant in level, is feedback controlled to allowappropriate slippage of the 3-4 clutch 53, so as to bring the turbinespeed change rate dNt into agreement with the target-turbine speedchange rate dNt₀. This shelf pressure control is accomplished bycontrolling not the second duty solenoid valve 122 generating the 3-4clutch pressure but the servo apply pressure through the first dutysolenoid valve 121. Specifically, as shown in FIG. 3, because thehydraulic pressure control circuit 100 is provided with an orifice 151at the junction where the hydraulic pressure line 224 extending from thesecond duty solenoid valve 122 is in communication with both hydraulicpressure line 225 and hydraulic pressure line 226 respectively leadingto the servo release pressure line 221 and the 3-4 clutch pressure line227, it is regarded that these hydraulic pressure lines 221 and 227 aredisconnected from the second duty solenoid valve 122 upstream therefromin terms of hydraulics. On the other hand, while the piston 54e of the2-4 brake 54 (see FIG. 4) is traveling in the cylinder 54d of the 2-4brake 54 by means of supply of the servo release pressure to the servorelease pressure chamber 54b of the 2-4 brake 54, it becomes hard tocontrol the hydraulic pressure in the pressure chamber of the 3-4 clutch53. Though, as shown in FIG. 4, because of the mechanical structure ofthe 2-4 brake 54 in which the piston 54e partitions the pressure chamberinto the servo apply pressure chamber 54a and the servo release pressurechamber 54b, the hydraulic pressure in the servo release pressurechamber 54b is directly affected by the hydraulic pressure in the servoapply pressure chamber 54a, enabling the hydraulic pressure in thepressure chamber of the 3-4 clutch 53 in communication with the servorelease pressure chamber 54b of the 2-4 brake 54, namely the 3-4 clutchpressure, to be controlled by means of controlling the servo applypressure through the first duty solenoid valve 121. Further, the secondduty solenoid valve 122 regulates the flowing quantity of working fluidsupplied to both pressure chamber of the 3-4 clutch 53 and servo releasepressure chamber 54b of the 2-4 brake 54 through the orifice 151. Thisregulation controls the passage of time for which the shelf pressure isheld in the inertia phase when the 3-4 clutch 53 is locked.

Accordingly, during the 2-3 up shift, while the level of shelf pressureis controlled by means of the first duty solenoid valve 121 when the 3-4clutch 53 is locked, the passage of time for which the shelf pressure isheld is controlled by means of the second duty solenoid valve 122.

Control of the first and second duty solenoid valves 121 and 122 isaccomplished following the flowcharts illustrating the sequence routinesshown in FIGS. 35 and 37, respectively. In the sequence routine of thefirst duty solenoid valve control shown by the flowchart in FIG. 35,steps S71 through S74 are identical in function with steps S1 through S4of the flowchart illustrating the first duty solenoid valve controlduring the 1-2 up shift shown in FIG. 14, respectively. When thecalculated hydraulic pressure Ps is obtained, after reading a lowerlimit hydraulic pressure Pg in accordance with the turbine torque Trfrom a limit pressure map such as shown in FIG. 36 at step S75, adetermination is made at step S76 as to whether or not the 2-3 up shifthas been concluded. Until a conclusion of the 2-3 up shift, thecalculated hydraulic pressure Ps is compared with the lower limithydraulic pressure Pg at step S77. When it is determined that thecalculated hydraulic pressure Ps is higher than the lower limithydraulic pressure Pg, the first duty solenoid valve 121 is operated ata duty rate corresponding to the calculated hydraulic pressure Ps atstep S78. On the other hand, when the calculated hydraulic pressure Psis equal to or lower than the lower limit hydraulic pressure Pg, thefirst duty solenoid valve 121 is operated at a duty rate correspondingto the lower limit hydraulic pressure Pg at step S79. In this instance,before the 2-3 up shift, the first duty solenoid valve 121 is operatedat a duty rate of 0% to supply a servo apply pressure, the prechargeperiod control is not performed. When it is determined at step S76 thatthe 2-3 up shift has concluded, the duty rate of the first duty solenoidvalve 121 is reduced at a fixed rate to 0% through steps S80 and S81,and the first duty solenoid valve control is terminated. As shown inFIG. 38, through the first duty solenoid valve control, the servo applypressure drops from a specified level to a level of shelf pressure, andincreases again to the specified level after the passage of time of theshelf pressure. While the shelf pressure remains held, both 3-4 clutchpressure and servo release pressure are controlled to hold a shelfpressure level corresponding to the servo apply pressure as labeled "E"in FIG. 38. As will be described in detail later, in cases where thetarget of servo apply pressure drops lower than the lower limithydraulic pressure Pg immediately after an appearance of a gear shiftcommand, the servo apply pressure is set to be equal to the lower limithydraulic pressure Pg as labeled "F."

In the sequence routine of the second duty solenoid valve control shownby the flowchart in FIG. 37 for the 2-3 up shift, the flowchart logiccommences and control passes directly to a function block at S91 whereinitialization is made to reset the timer count Tr to a specifiedinitial count Tr₀. Subsequently, after changing the timer count Tr by adecrement of 1 (one) at step S92, a determination is made at step S93 asto whether or not the precharge flag Fp has been up or whether or notthe timer count Tr is greater than 0 (zero) i.e. the timer count Tr hasbeen counted down smaller than the initial count Tr₀. When the prechargeflag Fp is up or when the timer count Tr has not yet counted downsmaller than the initial count Tr₀, the second duty solenoid valve 122is operated at a duty rate of 0% at step S94 to precharge rapidly thefluid paths leading to the pressure chamber of the 3-4 clutch 53 and theservo release pressure chamber 54b of the 2-4 brake 54, respectively,with the working fluid.

Thereafter, while the timer count Tr shows it counts of 0 (zero), i.e.has counted down over the initial count Tr₀, when the precharge flag Fpis reset to the state of "0," this indicates that the precharge periodhas terminated, then, at step S95, the second duty solenoid valve 122 isoperated at the same duty rate as for the first duty solenoid valve 121in the event where, while the 2-3 up shift is not concluded, thecalculated hydraulic pressure Ps is higher than the lower limithydraulic pressure Pg. As a result, the flowing quantities of workingfluid supplied to the pressure chamber of the 3-4 clutch 53 and theservo release pressure chamber 54b of the 2-4 brake 54 through theorifice 151 is reduced after the precharge period lower than during theprecharge period, and regulated to an appropriate quantity. The reasonwhy the termination of precharge period is determined on the basis notonly of the state of the precharge flag Fp but also of the passage oftime from an appearance of a gear shift command.

In the event where the first and second duty solenoid valves 121 and 122are operated at a same duty rate to supply a same level of workingpressure to the 2-4 brake 54 at the servo apply pressure chamber 54a andthe servo release pressure chamber 54b, because the piston 54e hasapproximately equal pressure surface areas on the sides of the servoapply pressure chamber 54a and the servo release pressure chamber 54b,the piston 54e is forced to travel in such a direction as to release the2-4 brake 54 only by the force of the spring 54i, and lays a relativelylong time on this travel. Thereafter, the control of second dutysolenoid valve 122 is terminated following the determination that thecontrol of first duty solenoid valve 121 has been over at step S96. As aresult, during the locking of the 3-4 clutch 53, the shelf pressure isheld for a duration of time sufficiently long for the inertia phase toconclude, eliminating a serious shift shock caused due, for example, toa sharp increase in working pressure resulting from a termination of theduration of shelf pressure before the conclusion of the inertia phase.

(8) Hydraulic Pressure Drop Control During 2-3 Up Shift

During the 2-3 up shift in which the 3-4 clutch pressure is indirectlycontrolled by means of the servo apply pressure during the 2-3 up shift,a servo apply pressure having been supplied at a relatively highpressure level in the state of second (2nd) gear is lowered once upon anappearance of a gear shift command and kept at the lowered pressurelevel for a duration of time sufficient to provide a torque phase inorder to initiate smooth control of the 3-4 clutch pressure in aninertia phase. In this case, the initial base pressure Pb', a correctionwith the learned pressure Pad of which is equivalent to the calculatedhydraulic pressure Ps in the torque phase and on the basis of which thefeedback control of servo apply pressure is initiated, is given througha correction of the hydraulic pressure Pi for a target turbine speedchange rate dNt₀ with hydraulic pressures Pt and Pt2 for target turbinetorque Tt₀ and its square (Tt₀)², respectively, as was described on a1-2 up shift. Accordingly, for example, when a 2-3 up shift isintentionally caused in a range of Row engine speeds by the driver, theinitial base pressure Pb', i.e. the servo apply pressure during a torquephase, is relatively low because the hydraulic pressure Pi is low.However, while the gear shift occurs, when the engine throttle openslargely and the engine encounters higher loads, the servo apply pressureis insufficient against input torque to the 2-4 brake 54, allowingslippage of the 2-4 brake 54 before commencement of the locking of the3-4 clutch 53 in the inertia phase which always results in an occurrenceof an engine blow.

In order to prevent the engine from encountering an engine blow duringthis kind of gear shift, at steps S95 through S99 of the flowchart inFIG. 35 the lower limit hydraulic pressure Pg set in accordance with theturbine torque Tt input to the transmission gear mechanism issubstituted for the calculated hydraulic pressure Ps for thedetermination of a duty rate at which the first duty solenoid valve 121is operated when it is exceeded by the calculated hydraulic pressure Ps.As a result, the servo apply pressure in the torque phase is preventedfrom falling below a pressure level corresponding to the input torque tothe 2-4 brake 54, arresting slippage of the 2-4 brake 54 beforecommencement of the locking of the 3-4 clutch 53 in the inertia phase.In the case where the servo apply pressure is fallen upon an appearanceof a gear shift command, while it is preferred for smooth commencementof the feedback control at the shift from a torque phase to an inertiaphase to make the servo apply pressure equal to the standard pressure onthe basis of which the feedback control is performed in the inertiaphase, nevertheless, the servo apply pressure may be fallen to anintermediate pressure level between the pressures before the gear shiftand during the inertia phase as labeled "G" in FIG. 39, or otherwise,may be gradually fallen from the pressure before the gear shift untilthe commencement of the inertia phase as labeled "H" in FIG. 40. In anycase, when the servo apply pressure is on the point of falling below thelower limit hydraulic pressure Pg set in accordance with the turbinetorque Tt in the torque phase, it is kept not to fall exceeding thelower limit hydraulic pressure Pg, preventing an occurrence of an engineblow.

(9) Precharge Control During 2-3 Up Shift

The precharge control during the 2-3 up shift is different from theprecharge control for the servo apply pressure chamber 54a of the 2-4brake 54 during the 1-2 up shift in respect that the precharge periodfor the servo apply and servo release pressure chambers 54a and 54b ofthe 2-4 brake 54 is determined on the basis not only of the state of theprecharge flag Fp but also of the passage of time from an appearance ofa gear shift command. As shown in FIG. 41, in general, when a workingpressure is supplied to a friction coupling element, such as the 3-4clutch 53, during up shifts, the working pressure increases as labeled"I" through a travel of the piston to an extreme end to initiate atorque phase. At this time, a torque draw such as labeled "J," whichrefers to a temporary drop in output torque To, occurs and is one ofcauses from which a serious shift shock results when input torque to thetransmission gear mechanism is large. In regard to making the adverseeffect of a torque draw to a gear shift feeling less strong, it can beconsidered to be effective to shorten the duration of torque draw or theduration of torque phase in which the torque draw occurs, or otherwiseto increase a rate at which the working pressure increases in the torquephase to eventually shorten the duration of torque phase. On account ofthis consideration, the precharge control is adequately performed toshorten the duration of torque phase.

As was previously described, as a result of the determination concerningthe up state of precharge flag Fp and the timer count Tr made at step403 of the flowchart shown in FIG. 37, while the precharge flag Fp isdown, when the timer count Tr is greater than 0 (zero), this indicatesthat the initial count Tr₀ has been counted down from an appearance of agear shift command, after terminating the precharge control for thefirst time, the control proceeds to the supply of calculated pressure Psthrough the second duty solenoid valve 122. In this instance, an initialcount map specifies the initial count Tr₀ so as to be greater as thethrottle opening θ becomes greater as shown in FIG. 42. Consequently,when input torque from the engine is great, the precharge control iscontinued after a shift of control to the inertia phase resulting fromhaving filled the specific fluid path with the working fluid, forexample as labeled "K", until the point of time of entering into theinertia phase as labeled "A" in FIG. 41. Resultingly, an increase in therate at which the working pressure rises in the torque phase as labeled"I" in FIG. 41 is yielded, which is accompanied by a shortened durationof torque phase or a torque draw occurring in the shortened duration oftorque phase, making aggravation of a gear shift feeling due to a shiftshock less intense. Because, during the 2-3 up shift, the shelf pressurecontrol on locking the 3-4 brake 53 is performed by controlling theservo apply pressure to the servo apply pressure chamber 54a of the 2-4brake 54 through the first duty solenoid valve 121, even continuing theprecharge control until the point of time of entering into the inertiaphase does not affect the control of inertia phase in which the turbinespeed Nt is fallen keeping the turbine speed change rate dNt inagreement with the target turbine speed change rate dNt₀. Consequently,while the control of inertia phase is satisfactorily maintained, a shiftshock during the torque phase is made less strong.

Because the initial count Tr₀ is set to 0 (zero) for throttle openings θless than a specified opening θ₀ as shown in FIG. 42, the determinationof termination of the precharge control is made on the basis of thestate of the precharge flag Fp only in a range of low input torque,during the 2-3 up shift similarly to during other up shifts.Consequently, during a gear shift in which the output torque To changesgently without being accompanied by a significant torque draw and asignificant torque blow, which refers to an phenomenon of a temporarysharp rise in output torque occurring at a shift of control from thetorque phase to the inertia phase immediately after the torque draw, asshown in FIG. 43, the condition for the gentle change of output torqueTo is maintained, preventing an occurrence of another shift shock due toincreasing the rate of rise of the working pressure.

The precharge control may be started at, instead of an appearance of agear shift command, the beginning of a torque phase at which the outputtorque To begins to fall for example. The precharge period may be setconstant. In this case, as shown in FIG. 44, the flowing quantity ofworking fluid passing through the second duty solenoid valve 122 isincreased larger with an increase in input torque by operating thesecond duty solenoid valve 122 at a duty rate of 0% for larger inputtorque and at a duty rate larger than 0% for smaller input torque.Furthermore, both precharge period and flowing quantity during theprecharge period may be changed in accordance with input torque.

It is to be understood that although the present invention has beendescribed with regard to preferred embodiments thereof, various otherembodiments and variants may occur to those skilled in the art, whichare within the scope and spirit of the invention, and such otherembodiments and variants are intended to be covered by the followingclaims.

What is claimed is:
 1. An automatic transmission control system forcontrolling an automatic transmission installed between an engine anddrive wheels of an automotive vehicle to multiply and transmit engineoutput to said drive wheels from said engine, said automatictransmission comprising a working fluid source operative to provideworking fluid pressure, a transmission gear mechanism with a pluralityof friction coupling elements including at least a first frictioncoupling element which has a locking pressure chamber and an unlockingpressure chamber divided by a piston and which is locked when only saidlocking pressure chamber of said first friction coupling element issupplied with said working fluid pressure and is unlocked when at leastsaid unlocking pressure chamber of said first friction coupling elementis supplied with said working pressure and when neither said lockingpressure chamber of said first friction coupling element nor saidunlocking pressure chamber of said first friction coupling element issupplied with said working fluid pressure and a second friction couplingelement which has a pressure chamber communicated with said unlockingpressure chamber of said first friction coupling element through acommunication fluid path and which is locked when said pressure chamberof said second friction coupling element is supplied with said workingfluid pressure, and a hydraulic pressure control system operative tocontrol supply of said working fluid pressure to and discharge of saidworking fluid pressure from pressure chambers of said friction couplingelements to selectively lock and unlock said friction coupling elementsaccording to driving conditions so as to change a torque transmissionpath in said transmission gear mechanism and thereby to provideavailable gears including a first specific gear which is achieved whensaid locking pressure chamber of said first friction coupling element issupplied with said working fluid pressure and neither said unlockingpressure chamber of said first friction coupling element nor saidpressure chamber of said second friction coupling element is suppliedwith said working fluid pressure and a second specific gear which isachieved when both said locking pressure chamber and said unlockingpressure chamber of said first friction coupling element and saidpressure chamber of said second friction coupling element are suppliedwith said working fluid pressure, said automatic transmission controlsystem comprising:a speed sensor operative to detect a vehicle speed ofthe automotive vehicle; a throttle position sensor operative to detect aposition of a throttle of said engine; a torque sensor operative todetect input torque transmitted to said transmission gear mechanism fromsaid engine; first working fluid pressure control means installed in afluid path extending from said working fluid source to both saidunlocking pressure chamber of said first friction coupling element andsaid pressure chamber of said second friction coupling element upstreamfrom said communication fluid path for controlling supply of saidworking fluid pressure to and discharge of said working fluid pressurefrom both said unlocking pressure chamber of said first frictioncoupling element and said pressure chamber of said second frictioncoupling element; second working fluid pressure control means installedin a fluid path extending from said fluid pressure source to saidlocking pressure chamber of said first friction coupling element forcontrolling supply of said working fluid pressure to and discharge ofsaid working fluid pressure from said locking pressure chamber of saidfirst friction coupling element; and control means for providing a gearshift command signal indicating one of said available gears selectedaccording to said throttle position and controlling said first workingfluid pressure control means and said second working fluid pressurecontrol means according to said gear shift command signal, said controlmeans controlling, during a specific gear shift from said first specificgear to said second specific gear responding to an appearance of aspecific gear shift command signal, said second working fluid pressurecontrol means according to said input torque such that a level of saidworking fluid pressure supplied to said locking pressure chamber of saidfirst friction coupling element for a substantial gear shift period oftime for which a change in input speed to said transmission gearmechanism occurs due to said specific gear shift becomes lower than alevel of said working fluid pressure supplied to said locking pressurechamber of said first friction coupling element immediately before saidappearance of said specific gear shift command signal and a level ofsaid working fluid pressure supplied to said locking pressure chamber ofsaid first friction coupling element after said substantial gear shiftperiod of time, and that, until a start of said substantial gear shiftperiod of time from said appearance of said specific gear shift commandsignal, a level of said working fluid pressure supplied to said lockingpressure chamber of said first friction coupling element becomes closeto a level of said working fluid pressure supplied to said lockingpressure chamber of said first friction coupling element during saidsubstantial gear shift period of time and retains said level of saidworking fluid pressure supplied to said locking pressure chamber of saidfirst friction coupling element higher than a specified level at whichsaid input speed to said transmission gear mechanism is allowed to runidle.
 2. The automatic transmission control system as defined in claim1, wherein said control means controls said second working fluidpressure control means such that, until a start of said substantial gearshift period of time from said appearance of said specific gear shiftcommand signal, said level of said working fluid pressure supplied tosaid locking pressure chamber of said first friction coupling elementbecomes equal to said level of said working fluid pressure supplied tosaid locking pressure chamber of said first friction coupling elementduring said substantial gear shift period of time.
 3. The automatictransmission control system as defined in claim 2, wherein said controlmeans controls said second working fluid pressure control means suchthat, until a start of said substantial gear shift period of time fromsaid appearance of said specific gear shift command signal, said levelof said working fluid pressure supplied to said locking pressure chamberof said first friction coupling element retains higher than a lowerlimit of said working fluid pressure specified according to said inputtorque so as to retain said level of said working fluid pressuresupplied to said locking pressure chamber of said first frictioncoupling element higher than said specified level.
 4. The automatictransmission control system as defined in claim 3, wherein said controlmeans controls said lower limit of said working fluid pressure to becomehigher with an increase in said input torque.
 5. An automatictransmission control system for controlling an automatic transmissioninstalled between an engine and drive wheels of an automotive vehicle tomultiply and transmit engine output to said drive wheels from saidengine, said automatic transmission comprising a working fluid sourceoperative to provide working fluid pressure, a transmission gearmechanism with a plurality of friction coupling elements including atleast a first friction coupling element which has a locking pressurechamber and an unlocking pressure chamber divided by a piston and whichis locked when only said locking pressure chamber of said first frictioncoupling element is supplied with said working fluid pressure and isunlocked when at least said unlocking pressure chamber of said firstfriction coupling element is supplied with said working pressure andwhen neither said locking pressure chamber of said first frictioncoupling element nor said unlocking pressure chamber of said firstfriction coupling element is supplied with said working fluid pressureand a second friction coupling element which has a pressure chambercommunicated with said unlocking pressure chamber of said first frictioncoupling element through a communication fluid path and which is lockedwhen said pressure chamber of said second friction coupling element issupplied with said working fluid pressure, and a hydraulic pressurecontrol system operative to control supply of said working fluidpressure to and discharge of said working fluid pressure from pressurechambers of said friction coupling elements to selectively lock andunlock said friction coupling elements according to driving conditionsso as to change a torque transmission path in said transmission gearmechanism and thereby to provide available gears including a firstspecific gear which is achieved when said locking pressure chamber ofsaid first friction coupling element is supplied with said working fluidpressure and neither said unlocking pressure chamber of said firstfriction coupling element nor said pressure chamber of said secondfriction coupling element is supplied with said working fluid pressureand a second specific gear which is achieved when both said lockingpressure chamber and said unlocking pressure chamber of said firstfriction coupling element and said pressure chamber of said secondfriction coupling element are supplied with said working fluid pressure,said automatic transmission control system comprising:a speed sensoroperative to detect a vehicle speed of the automotive vehicle; athrottle position sensor operative to detect a position of a throttle ofsaid engine; a torque sensor operative to detect input torquetransmitted to said transmission gear mechanism from said engine; afirst solenoid valve installed in a fluid path extending from saidworking fluid source to both said unlocking pressure chamber of saidfirst friction coupling element and said pressure chamber of said secondfriction coupling element upstream from said communication fluid pathfor controlling supply of said working fluid pressure to and dischargeof said working fluid pressure from both said unlocking pressure chamberof said first friction coupling element and said pressure chamber ofsaid second friction coupling element; a second solenoid valve installedin a fluid path, which extends from said fluid pressure source to saidlocking pressure chamber of said first friction coupling element, forcontrolling supply of said working fluid pressure to and discharge ofsaid working fluid pressure from said locking pressure chamber of saidfirst friction coupling element; and a controller operative to provide agear shift command signal indicating one of said available gearsselected according to said vehicle speed, said throttle position andsaid input torque and to control said first and second solenoid valvesaccording to said gear shift command signal, said controller, during aspecific gear shift from said first specific gear to said secondspecific gear responding to an appearance of a specific gear shiftcommand signal, said second solenoid valve according to said inputtorque such that a level of said working fluid pressure supplied to saidlocking pressure chamber of said first friction coupling element for asubstantial gear shift period of time for which a change in input speedto said transmission gear mechanism occurs due to said specific gearshift becomes lower than a level of said working fluid pressure suppliedto said locking pressure chamber of said first friction coupling elementimmediately before said appearance of said specific gear shift commandsignal and a level of said working fluid pressure supplied to saidlocking pressure chamber of said first friction coupling element aftersaid substantial gear shift period of time, and that, until a start ofsaid substantial gear shift period of time from said appearance of saidspecific gear shift command signal, a level of said working fluidpressure supplied to said locking pressure chamber of said firstfriction coupling element becomes close to a level of said working fluidpressure supplied to said locking pressure chamber of said firstfriction coupling element during said substantial gear shift period oftime and retains said level of said working fluid pressure supplied tosaid locking pressure chamber of said first friction coupling elementhigher than a specified level at which said input speed to saidtransmission gear mechanism allows to run idle.
 6. The automatictransmission control system as defined in claim 5, wherein saidcontroller controls said second solenoid valve such that, until a startof said substantial gear shift period of time from said appearance ofsaid specific gear shift command signal, said level of said workingfluid pressure supplied to said locking pressure chamber of said firstfriction coupling element retains higher than a lower limit of saidworking fluid pressure specified according to said input torque so as toretain said level of said working fluid pressure supplied to saidlocking pressure chamber of said first friction coupling element higherthan said specified level.
 7. The automatic transmission control systemas defined in claim 6, wherein said controller controls said lower limitof said working fluid pressure to become higher with an increase in saidinput torque.